Generally, hydraulic construction machines such as power shovels and cranes are driven by a hydraulic motor or actuator in travelling or turning operations. For the purpose of preventing imposition of excessive inertial loads on the hydraulic motor at the time of stopping a travelling or turning operation, a brake valve is usually inserted in the hydraulic drive circuit thereby to absorb the inertial loads through conversion into thermal energy of the operating hydraulic pressure.
Such a brake valve is constituted, for example, by a counterbalance valve which is provided between a hydraulic pressure source and an actuator, a couple of relief valves located in positions closer to the actuator than to the counterbalance valve, and a couple of check valves located between the counterbalance valve and the respective relief valves (e.g., Japanese Utility Model Publication 62-31681).
The just-mentioned prior art brake valve is shown in FIGS. 6 through 9.
In these figures, indicated at 1 is a hydraulic motor serving as an actuator and constructed in the form of a radial piston type hydraulic motor, from a control pin 8, a rotor 11 and a piston 14, which will be described hereinlater. The hydraulic motor 1 has its inlet/outlet ports connected to a hydraulic pump 3 or a hydraulic pressure source through a pair of conduits 2A and 2B, so that it is rotationally driven by the operating oil pressure discharged from the hydraulic pump 3 which sucks in oil from a tank 4.
Indicated at 5 is a direction change-over valve which is located within the lengths of the conduits 2A and 2B between the hydraulic motor 1 and pump 3. The direction change-over valve 5 is manipulated by an operator to and from a neutral position (a) where the hydraulic motor 1 is at rest and two change-over positions (b) or (c) where the motor 1 is rotated either in the direction of arrow F or R.
The reference numeral 6 denotes a brake valve which is located within the lengths of the conduits 2A and 2B between the hydraulic motor 1 and direction change-over valve 5, and which is largely constituted by counterbalance valve 23, relief valves 28A and 28B, and check valves 39A and 29B, which will be described later. When the direction change-over valve 5 is held in the neutral position (a), the brake valve 6 opens either the relief valve 28A or 28B to relieve the oil pressure to the conduit 2A or 2B whichever is on the lower pressure side, thereby to apply brakes to the hydraulic motor 1.
FIG. 7 illustrates the hydraulic motor 1 in greater detail, in which the reference 7 indicates a motor casing of a lidded cylindrical shape having a cam surface 7A formed completely around its entire inner periphery and having its opening lidded with a valve casing 18.
Denoted at 8 is a control pin which is formed integrally with the valve casing 18 in such a manner as to project substantially from a center portion of the latter, the axis of the control pin 8 being offset from the axis of the cam surface 7A to a predetermined extent. In an axially intermediate portion, the control pin 8 is formed with a pair of charging/discharging passages 9 and 10, which are opened on the outer peripheral surface of the control pin 8 to serve as inlet/outlet ports for the hydraulic motor 1, as seen also in FIG. 8.
Indicated at 11 is a rotor which is rotatably mounted on the control pin 8 and which is provided with a plural number of radially extending cylinders 11A (only one of which cylinders 11A is shown in the drawing) at angularly spaced positions around the circumference thereof. Each one of the cylinders 11A is intermittently communicated with the charging/discharging passages 9 and 10 through a port 11B. The axis of the rotor 11 is disposed eccentrically relative to the axis of the cam surface 7A, forming an eccentric space 12 of a crescent shape between its outer periphery and the cam surface 7A.
Designated at 13 is an output shaft which is integrally provided on one end face of the rotor 11. This output shaft 13 is coupled with an external inertial body through a reducer (not shown) or the like, and rotatable integrally with the rotor 11 to transmit the rotation of the latter to the outside.
Indicated at 14 are pistons which are reciprocably received in the cylinders 11A, at 15 are balls which are rockably provided in the respective pistons 14, and at 16 are shoes which are located between the balls 15 and the cam surface 7A and in spaced positions in the circumferential direction. Each shoe 16 is fitted in an opposing ball 15 at its fore end and slidably engaged with the cam surface 7A at its rear end through guides 17.
Referring to FIGS. 7 to 9, the description is now directed to the brake valve 6.
As seen in these figures, the afore-mentioned valve casing 18 is attached to the open side of the motor casing 7 as a closure lid and firmly fixed to the latter by means of a plural number of bolts 19. The valve casing 18 is integrally formed with oil passages 20A and 20B, as will be described later, and provided with a counterbalance valve 23, relief valves 28A and 28B, and check valves 39 and 39B.
The oil passages 20A and 20B which are formed internally of the valve casing 18 constitute part of the conduits 2A and 2B, respectively. The oil passages 20A and 20B consist of oil passages 21A and 21B, which are located on the side of the oil pressure source and connected to the hydraulic pump 3 through the direction change-over valve 5, and oil passages 22A and 22B, which are located on the side of the actuator and connected to the hydraulic motor 1 through the charging/discharging passages 9 and 10, respectively. Through check valves 39A and 39B, the oil passages 21A and 21B on the side of the hydraulic pressure source are communicated with the oil passages 22A and 22B on the side of the actuator, respectively.
Indicated at 23 is the counterbalance valve which is provided in the valve casing 18 in a position closer to the hydraulic pump 3. The counterbalance valve 23 is largely constituted by a spool sliding bore 24 which is integrally formed in the valve casing 18, and a spool 25 which is slidably fitted in the spool sliding bore 24. The spool 25 is provided with a land 25A which establishes or blocks communication between the oil passages 21A and 22A, and a land 25B which establishes or blocks communication between the oil passages 21B and 22B. The spool 25 has its opposite ends disposed in oil chambers 26A and 26B, respectively, and is urged into a neutral position by return springs 27A and 27B in the oil chambers 26A and 26B. The above-mentioned counterbalance valve 23 is operated in interlinked relation with the direction change-over valve 5, and switchable to either the change-over position (b) or (c) from the neutral position (a).
The references 28A and 28B denote a pair of relief valves which are provided in the valve casing 18 in positions closer to the hydraulic motor 1. As shown also in FIG. 9, the relief valves 28A and 28B include valve guides 29A and 29B, main valve bodies 32A and 32B and pistons 36A and 36B, respectively, to constitute crossover relief valves with the so-called shockless function, as will be described later.
The valve guides 29A and 29B, which constitute part of the relief valves 28A and 28B, are provided with passage holes 30A and 30B at the respective fore ends, which are disposed in the oil passages 22A and 22B on the side of the actuator.
The references 31A and 31B denote valve seat members of cylindrical shape, which are located opposingly to the valve guides 29A and 29B, and are in communication with the oil passages 22A and 22B on the side of the actuator, respectively. The main valve bodies 32A and 32B are slidably received in the valve guides 29A and 29B for seating on or unseating off the valve seat members 31A and 31B, respectively. The main valve bodies 32A and 32B are constantly urged in the closing direction by valve springs 33A and 33B, and provided with axial throttle passages 34A and 34B, respectively.
In this instance, the main valve bodies 32A and 32B are so dimensioned as to hold the relationship of d.sub.&gt;d.sub.2 where d.sub.1 is the diameter of the seating portions to be engaged with the valve seat members 29A and 29B and d .sub.2 is the diameter of the sliding portions in the valve guides 29A and 29B.
In case the pressure P1 in the respective valve seat members 31A and 31B is equal with the pressure P2 in the respective valve guides 29A and 29B, the main valve bodies 32A and 32B are disengaged from the valve seat members 31A and 31B against the action of the valve springs 33A and 33B to go into a high pressure relief action as soon as the pressure P1 reaches a predetermined valve opening pressure P0, due to the difference in pressure receiving area between d.sub.1 and d.sub.2.
On the other hand, in case the pressure P2 in the valve guides 29A and 29B is maintained at a level lower than the pressure P1 in the valve seat members 31A and 31B by the operation of floating pistons 38A and 38B, which will be described later, even if the pressure P1 is at a level lower than the predetermined valve opening pressure P0, the main valve bodies 32A and 32B are disengaged from the valve seat members 31A and 31B against the action of the valve springs 33A and 33B, respectively, communicating the oil passages 22A and 22B with each other to effect a low pressure relief action.
The references 35A and 35B indicate annular oil chambers which are formed between the valve casing 18 and the outer peripheries of the valve guides 29A and 29B, the inner ends of the oil chambers 35A and 35B being communicated with the oil passages 22A and 22B on the side of the actuator, respectively. The references 36A and 36B indicate annular pistons which are slidably received in the oil chambers 35A and 35B, the pistons 36A and 36B forming floating pistons 38A and 38B, as shown also in FIG. 6, in cooperation with throttle passages 37A and 37B axially formed in the valve guides 29A and 29B and the oil chambers 35A and 35B, respectively. As the oil pressure in the passages 22A and 22B on the side of the actuator flows into the valve guides 29A and 29B through the throttle passages 34A and 34B and then into the spaces at the outer ends of the oil chambers 35A and 35B through the throttle passages 37A and 37B, the pistons 36A and 36B are moved toward the fore ends of the valve guides 29A and 29B until they abut against the valve casing 18. At this time, the relief valves 28A and 28B maintain the valve opening pressure at a low level. Namely, the time period of displacement of the pistons 36A and 36B corresponds to the low pressure relief time.
Indicated at 39A and 39B are a pair of check valves which are provided within the lengths of the oil passages 20A and 20B at positions between the counterbalance valve 23 and the respective relief valve 28A or 28B. By the actions of valve springs 40A and 40B, these check valves 39A and 39B are constantly urged in the valve closing direction to seat on valve seats 41A and 41B which are formed between the oil passages 21A and 21B on the side of the pressure source and the oil passages 22A and 22B on the side of the actuator, respectively. Further, when the oil pressure from the hydraulic pump 3 is introduced into the oil passages 21A and 21B on the side of the pressure source, the check valves 39A and 39B are opened by the oil pressure against the action of the valve springs 40A and 40B, respectively, thereby permitting the oil pressure to flow into the oil passages 22A and 22B on the side of the actuator while blocking reverse flows of the oil pressure.
The reference numeral 42 denotes a shuttle valve which is provided in the valve casing 18 at a position closer to the hydraulic pump 3 than to the counterbalance valve 23 and which is in communication with the oil passages 21A and 21B on the side of the pressure source. The shuttle valve 42 selects either the oil passage 21A or 21B whichever is at a higher pressure level, for supplying part of the oil pressure to a brake device (not shown) or the like as a pilot pressure.
The prior art brake valve with the above-described construction operates in the manner as follows.
Firstly, if the direction change-over valve 5 is switched by an operator to the change-over position (b) from the neutral position (a), the oil pressure which is discharged from the hydraulic pump 3 is allowed to flow into the oil passages 21A on the side of the pressure source through the conduit 2A. Then, due to a pressure differential between the oil passage 21A on the side of the pressure source and the oil passage 22A on the side of the actuator, the check valve 39A is opened against the action of the valve spring 40A, and the oil pressure in the oil passage 21A on the side of the pressure source is allowed to flow into the cylinders 11A of the rotor 11 through the oil passage 22A on the side of the actuator and the respective charging/discharging passages 9.
As a result, the pistons 14 are put in reciprocating movement within the cylinders 11A, causing the shoes 16 to slide along the cam surface 7A and turning the rotor 11 about the control pin 8. The rotation of the rotor 11 is led out through the output shaft 13 to drive the inertial body into rotation.
When the oil pressure is introduced into the oil passage 21A on the side of the pressure source, part of the oil is fed into the oil chamber 26A through the throttle passage of the spool 25, thereby urging the spool 25 in the rightward direction in FIG. 8. Consequently, the land 25B is moved to the right to communicate the oil passage 21B on the side of the pressure source with the oil passage 22B on the side of the actuator. That is to say, the counterbalance valve 23 is consequently switched to the change-over position (b) from the neutral position (a) shown in FIG. 6. The oil pressure which is pushed out of each cylinder 11A during its compression stroke is fed to the oil passage 22B on the side of the actuator through the charging/discharging passages 10, and then discharged from the oil passage 22B to the outside of the brake valve 6 through the oil passage 21B on the side of the pressure source, thereafter the oil being recirculated to the tank 4 through the direction change-over valve 5 and conduit 2.
Conversely, when the operator switches the direction change-over valve 5 from the change-over position (b) to the neutral position (a) to stop the rotation of the inertial body, the check valve 39A is urged to seat on the valve seat 41A by the action of the valve spring 40A, blocking communication between the oil passage 21A on the side of the pressure source and the passage 22A on the side of the actuator. Further, at the counterbalance valve 23, as a result of a pressure drop in the oil chamber 26A, the spool 25 is urged to return to the neutral position (a) by the biasing action of the return spring 27B, thereby blocking communication between the oil passage 21B on the side of the pressure source and the oil passage 22B on the side of the actuator.
However, even after the direction change-over valve 5 has been switched to the neutral position (a), the hydraulic motor 1 is forced to rotate continuedly under the influence of the inertial force of the inertial body which is coupled with the motor output shaft 13, still keeping the pumping action, sucking in the oil pressure from the primary oil passage 22A on the side of the actuator and discharging same to the secondary oil passage 22B on the side of the actuator. The pressure in the secondary oil passage 22B on the side of the actuator is gradually increased since return of oil pressure to the tank 4 is blocked by the counterbalance valve 3 and the check valve 39B.
Then, the oil pressure in the oil passage 22B on the side of the actuator flows into the valve guide 29B through the throttle passage 34B to act on the piston 36B from the valve guide 29B via throttle passage 37B. As a consequence, the piston 36B is slided within the oil chamber 35B toward the primary oil passage 22A on the side of the actuator. In the meantime, due to pressure losses in the throttle passages 34B and 37B, the pressure in the valve guide 29B is maintained at a lower level than the pressure in the oil passage 22B on the side of the actuator, so that the main valve body 32B is opened at a pressure level lower than the predetermined valve opening pressure Po for a relief at low pressure. Then, as the piston 36B is stopped by abutment against the valve casing 18, the pressure in the valve guide 29B becomes equal with the pressure in the oil passage 22B on the side of the actuator and rises to the preset valve opening pressure level to effect high pressure relief.
While the relief valve 28B is open, the oil passages 22A and 22B on the side of the actuator are communicated with each other, forming a closed circuit together with the charging/discharging passages 9 and 10. As a result, the oil pressure which is discharged from the hydraulic motor 1 is converted into thermal energy while being passed through the relief valve 28B, thereby absorbing the inertial force of the inertial body to produce a braking force.
In case the direction change-over valve 5 is switched to the change-over position (c) from the neutral position (a), the brake valve operates substantially in a similar manner, and therefore accounts in this regard are omitted to avoid repetitions.
In the above-described prior art brake valve employing the relief valves 28A and 28B in combination with accumulators 38A and 38B for the shockless function, the oil pressure is relieved at low pressure for a predetermined time period (for a predetermined time period of low pressure relief) until the pistons 36A and 36B in sliding movement are stopped, thereby preventing the shocks which might result from abrupt stoppage of the inertial body by sudden application of braking forces.
However, according to the prior art, even when the direction change-over valve is switched to the change-over position (b) or (c) from the neutral position (a) to drive the hydraulic motor 1, the oil pressure discharged from the hydraulic pump 3 acts on the pistons 36A and 36B through the oil passages 22A and 22B on the side of the actuator and the throttle passages 34A and 34B. Accordingly, the drive pressure is retained at a low pressure level until the pistons 36A and 36B are stopped by abutment against the valve casing 18. This involves a problem of low response characteristics such as a delay in driving the hydraulic motor 1 under high load conditions, for example, in a hill climbing or steering operation.
Besides, the hydraulic motor 1 is driven abruptly upon lapse of a predetermined time period (a time period of low pressure relief) or when the pistons 36A and 36B in sliding movement come to a stand, so that the operator feels as if the inertial body were suddenly put in operation. This can deteriorate the safety of operation to a considerable degree.
Moreover, the time period of low pressure relief by the accumulators 38A and 38B is determined by the flow areas of the throttle passages 34A and 34B, bored in the main valve bodies 32A and 32B, and of the throttle passages 37A and 37B. Therefore, further deteriorations in response characteristics take place at lower ambient temperatures which are reflected by a higher viscosity of the operating oil and a longer time period of low pressure relief as compared with normal or ordinary ambient temperatures. This also gives rise to a problem that the safety of operation is impaired to a considerable degree.
In view of the above-discussed problems of the prior art, the present invention has as its object the provision of a brake valve which can suppress the low pressure relief by relief valves at the time of starting an inertial body, which serves as an actuator, for the purpose of improving the response characteristics and safety of operation.